JP4092474B2 - COMPRESSION RATIO CONTROL DEVICE FOR INTERNAL COMBUSTION ENGINE - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a compression ratio control device for an internal combustion engine using a multi-link type piston-crank mechanism, and more particularly to improvement of an actuator portion that performs variable control of the compression ratio.
[0002]
[Prior art]
The present applicant previously used a multi-link type piston-crank mechanism as a variable compression ratio mechanism of a reciprocating internal combustion engine, and changed the piston top dead center position by moving a part of the link configuration. Various mechanisms have been proposed (for example, Patent Document 1). The upper link is connected to the piston via a piston pin, the lower link is connected to the upper link via the upper pin so as to be swingable, and is rotatably attached to the crank pin of the crankshaft. A control link that is swingably connected to the lower link via a control pin, and an eccentric cam portion that is rotatably provided to the cylinder block and that supports the other end of the control link so as to be swingable. The engine compression ratio is variably controlled by controlling the rotational position of the control shaft (that is, the eccentric direction of the eccentric cam portion) according to engine operating conditions via an actuator. It is the composition to do.
[0003]
[Patent Document 1]
Japanese Patent Laid-Open No. 2002-215902
[Problems to be solved by the invention]
In the configuration as described above, a load applied to the control link during operation of the engine acts on the eccentric cam portion, and as a result, a rotational moment is generated in the control shaft. Therefore, the actuator that drives the control shaft in the rotational direction is It is necessary to counter this rotational moment while maintaining the compression ratio constant as well as during the change of the compression ratio. Here, if the energy consumed for driving the actuator is excessively large, the fuel efficiency improvement effect by the variable control of the compression ratio is remarkably offset. On the other hand, there is a problem that knocking occurs if the compression ratio is not quickly changed to a low compression ratio, for example, during a transition that changes from a low load to a high load.
[0005]
The present invention takes into account the problems peculiar to the compression ratio control apparatus as described above, and can ensure the control response while minimizing the energy consumption accompanying the drive of the actuator. The purpose is to provide.
[0006]
[Means for Solving the Problems]
According to a first aspect of the present invention, there is provided a compression ratio control apparatus for an internal combustion engine, wherein the first link is connected to a piston reciprocating in a cylinder via a piston pin, and swings on the first link. A second link rotatably connected to the crank pin of the crankshaft, a third link having one end swingably connected to the second link, and a cylinder block rotatable. A control shaft provided with an eccentric cam portion that swingably supports the other end portion of the third link, and an actuator that controls the rotational position of the control shaft. The engine compression ratio is changed by controlling the rotational position.
[0007]
In particular, according to the present invention, when the compression ratio is controlled to the highest value, a plane passing through the center line of the control shaft and the center line of the eccentric cam portion extends from the third link to the eccentric cam portion. It is comprised so that it may become substantially parallel with respect to the load vector of the load which acts on. The direction of the load acting from the third link changes slightly with the stroke of the piston. In the present invention, the “load vector” means a vector of the average value of the loads. To do.
[0008]
If comprised in this way, a moment will not generate | occur | produce in a control shaft by the said load vector, while the compression ratio is controlled to the highest value. Alternatively, the moment is extremely small. Therefore, the energy required for maintaining the rotational position of the control shaft is very small. In addition, since the compression ratio has the highest value in the partial load region where low and medium speeds are frequently used, the ratio of operating time that does not consume wasteful energy is high, and thus energy consumption is effective. Can be suppressed.
[0009]
More preferably, even when the compression ratio is controlled to the lowest value, the plane is preferably configured to be substantially parallel to the load vector. This can be realized by setting the rotation angle range of the control shaft to approximately 180 °. The low compression ratio condition is generally a high load condition, and a load acting on the eccentric cam portion from the third link via the first and second links becomes large. Therefore, at this time, it is advantageous in terms of durability and reliability of the actuator by preventing the rotation moment from being generated in the control shaft. In particular, in the case of a hydraulic accumulator, the problem of leakage of hydraulic oil can be avoided. Furthermore, by setting the rotation angle range of the control shaft to approximately 180 °, the amount of eccentricity of the eccentric cam portion is minimized, and the rotational moment due to the load vector is suppressed accordingly even in the intermediate compression ratio state.
[0010]
In addition, when the compression ratio is controlled to the highest value, the plane is not completely parallel to the load vector, and a moment that causes the control shaft to rotate toward the low compression ratio is generated by the load vector. It can also be configured to. If comprised in this way, it will become advantageous on the responsiveness at the time of changing from a high compression ratio state to a low compression ratio at the time of a transition.
[0011]
【The invention's effect】
According to the present invention, since almost no rotational moment is generated in the control shaft at the highest compression ratio that is frequently used, the energy consumed by driving the actuator can be minimized without any loss in control response.
[0012]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings.
[0013]
FIG. 1 shows an embodiment of a compression ratio control apparatus for an internal combustion engine according to the present invention. This internal combustion engine is, for example, an in-line four-cylinder spark ignition gasoline engine, and is weak based on a detection signal from a variable compression ratio mechanism 1 that variably controls the nominal compression ratio ε and a knock sensor 3 that detects knocking. An ignition advance control device 2 that controls the ignition timing and an engine control unit 4 that controls the variable compression ratio mechanism 1 and the ignition advance control device 2 are provided so as to be in a knocking state. The engine control unit 4 is provided with a compression ratio control map in which a target compression ratio is assigned in advance corresponding to the engine operating conditions, and an engine speed signal, a load signal, a suction signal detected by sensors (not shown). Negative pressure signal, exhaust temperature signal, etc. are input. The variable compression ratio mechanism 1 is subjected to compression ratio control by a rotary hydraulic actuator 5 as will be described later. The engine control unit 4 drives the hydraulic actuator 5 via a hydraulic device 6. Yes.
[0014]
FIG. 2 is a diagram showing a configuration of the variable compression ratio mechanism 1.
[0015]
The crankshaft 51 includes a plurality of journal portions 52 and a crankpin portion 53, and the journal portion 52 is rotatably supported by the main bearing of the cylinder block 50. The crankpin portion 53 is eccentric from the journal portion 52 by a predetermined amount, and a lower link 54 serving as a second link is rotatably connected thereto.
[0016]
The lower link 54 is configured to be split into two left and right members, and the crank pin portion 53 is fitted in a substantially central connecting hole.
[0017]
The upper link 55 serving as the first link has a lower end side rotatably connected to one end of the lower link 54 by a connecting pin 56, and an upper end side rotatably connected to a piston 58 by a piston pin 57. The piston 58 receives the combustion pressure and reciprocates in the cylinder 59 of the cylinder block 50.
[0018]
The control link 60 serving as the third link is pivotally connected at its upper end side to the other end of the lower link 54 by a connecting pin 61, and the lower end side of the lower part of the cylinder block 50 that becomes part of the engine body via the control shaft 62. It is connected to the pivotable. Specifically, the control shaft 62 is rotatably supported by the engine body and has an eccentric cam portion 62a that is eccentric from the center of rotation, and the lower end portion of the control link 60 is rotatable on the eccentric cam portion 62a. Is fitted.
[0019]
As shown in FIG. 3, the control shaft 62 is arranged in parallel with the crankshaft 51 along the cylinder row direction, the eccentric cam portions 62a are formed at four locations, and the control links 60 of the four cylinders are provided. Each is linked. The hydraulic actuator 5 is coaxially attached at the end in the engine longitudinal direction, and the rotational position of the hydraulic actuator 5 is controlled by the actuator 5.
[0020]
In the variable compression ratio mechanism 1 using the multi-link type piston-crank mechanism as described above, when the control shaft 62 is rotated by the hydraulic actuator 5, the center position of the eccentric cam portion 62a, in particular, the engine body. The relative position with respect to changes. Thereby, the rocking | fluctuation support position of the lower end of the control link 60 changes. When the swing support position of the control link 60 changes, the stroke of the piston 58 changes, and the position of the piston 58 at the piston top dead center (TDC) becomes higher (see FIG. 7) or lower (see FIG. 8). refer. This makes it possible to change the engine compression ratio. 7 and 8 representatively show the high compression ratio state and the low compression ratio state, but the compression ratio can be continuously changed between them.
[0021]
FIG. 5 shows the compression ratio control characteristic of the variable compression ratio mechanism 1, in other words, the target compression ratio characteristic corresponding to the operating conditions set in the compression ratio control map. FIG. 5 shows an example of a supercharged engine. As shown in FIG. 5, generally, the engine is controlled so as to have a lower compression ratio as the engine load increases.
[0022]
Further, in the above-described multi-link variable compression ratio mechanism 1, a piston stroke characteristic close to simple vibration can be obtained by appropriately selecting a link dimension. This stroke characteristic close to simple vibration is advantageous in terms of vibration noise, but in particular, there is an advantage that the piston speed near the top dead center becomes gentler than that of a general single link type piston-crank mechanism.
[0023]
An example of a specific link configuration for bringing the piston motion close to a simple vibration motion is as follows.
[0024]
First, as shown in FIG. 4A, the rotation center of the crankshaft 51 (axial center of the journal portion 52) O is the origin, and the direction orthogonal to the piston pin 57 and its reciprocating axis 1 (thrust-anti-thrust direction) When the x axis is taken in parallel, the y axis is taken in parallel with the reciprocation axis 1 of the piston pin 57, and the rotation direction of the crankshaft 51 is defined as the counterclockwise direction, the reciprocation passes through the axis Oc of the piston pin 57. The x-coordinate of the axis (≈the axis of the cylinder 59) l is set to a negative value, and the x-coordinate of the axis Oa of the eccentric cam portion 62a serving as the swing center of the control link 60 is set to a positive value. .
[0025]
And as shown in FIG.
The distance between the axis O of the crankshaft 51 and the axis Oe of the crankpin 53 is L1;
The distance between the shaft center Oe of the crankpin portion 53 and the shaft center Of of the connecting pin 61 that connects the lower link 54 and the control link 60 so as to be relatively rotatable is L2;
The link length of the control link 60 is L3;
The distance between the shaft center Oe of the crank pin portion 53 and the shaft center Od of the connecting pin 56 that connects the upper link 55 and the lower link 54 so as to be relatively rotatable is L4;
The distance between the axis Of and the axis Od is L5;
The link length of the upper link 55 is L6;
The coordinate position of the swing axis Oa of the control link 60 is (XC, YC);
X coordinate of the reciprocal axis l of the piston pin 57 is x4;
Is defined so that the following ratio holds.
[0026]
[Expression 1]
L1: L2: L3: L4: L5: L6: XC: YC: x4
≈ 1: 2.4: 2.65 to 3.5: 0.69: 3.0 to 3.4: 3.3 to 3.55: 3.2 to 3.55: -2 to -1.35 : -1 to -0.6
XC and YC vary depending on the rotational position of the control shaft 62, but are set so that the above ratio is always established when the rotational position of the control shaft 62 is within the control range.
[0027]
With such a link configuration, the piston motion approximates a single vibration motion, and the piston motion near the top dead center becomes gentle.
[0028]
Next, FIG. 6 shows the characteristics of the moment load depending on the angle θ formed by the eccentric direction of the eccentric cam portion 62a and the load vector. The eccentric direction of the eccentric cam portion 62a is defined by a plane P that passes through the center line of the control shaft 62 and the center line of the eccentric cam portion 62a. As shown in the figure, the moment load is 0 when the angle θ is 0 °, takes a maximum value when θ is 90 °, decreases again when θ increases, and approaches 0. According to the compression ratio control map shown in FIG. 5, the maximum compression ratio is basically set in the partial load operation region where the low / medium speed is frequently used. It is possible to maintain the control state with minimum energy.
[0029]
FIG. 7 shows the posture of the eccentric cam portion 62a of the control shaft 62 and the direction of the load received by the control shaft 62 in a state where the compression ratio is the highest. In the above link configuration, the eccentric cam portion 62a is positioned relatively low, so that a high compression ratio is obtained. In this embodiment, however, the eccentric direction, that is, the center line of the control shaft 62 and the center line of the eccentric cam portion 62a. Are substantially parallel to the load vector, that is, substantially coincide with each other. In other words, the angle θ is substantially zero. As a result, the control shaft 62 does not receive a rotational moment due to a load acting on the eccentric cam portion 62a.
[0030]
FIG. 8 shows a state in which the compression ratio is controlled to the lowest, and the angle θ formed by the eccentric direction of the eccentric cam portion 62a and the load vector is 180 °. That is, in the above-described link configuration, the eccentric cam portion 62a is positioned relatively high, resulting in a low compression ratio. In this embodiment, the center line of the control shaft 62 and the center line of the eccentric cam portion 62a are The passing plane P is substantially parallel to the load vector, that is, substantially coincides with each other. Therefore, even in this control state, the control shaft 62 does not receive a rotational moment due to the load acting on the eccentric cam portion 62a.
[0031]
That is, in this embodiment, the rotation angle range of the control shaft 62 is approximately 180 °, and the moment load can be reduced under the conditions at both ends of the variable compression ratio range. Since the rotation angle of the control shaft 62 is large in this way, the eccentric amount of the eccentric cam portion 62a itself is a minimum of about 1/2 of the variable range required in the vertical direction. Therefore, the rotational moment generated by the load vector is relatively small at the intermediate rotational position.
[0032]
When the compression ratio is controlled to the highest value as shown in FIG. 7, the plane P is slightly inclined with respect to the load vector, and the control shaft 62 is moved to the low compression ratio side by the load vector. It can also be configured so as to generate a moment of rotation. If comprised in this way, it will become advantageous on the responsiveness at the time of changing from a high compression ratio state to a low compression ratio at the time of a transition. In this case, the rotation angle range of the control shaft 62 is slightly smaller than 180 °.
[0033]
FIG. 9 shows details of the hydraulic device 6 of FIG. 1 that drives the hydraulic actuator 5. As described above, the rotation angle range of the control shaft 62 is about 180 °. As the hydraulic actuator 5, for example, a single vane in which the first hydraulic chamber 12 and the second hydraulic chamber 13 are partitioned by one vane 11. A structure is used.
[0034]
The first and second hydraulic chambers 12 and 13 are appropriately connected to the hydraulic pressure introduction passage 16 and the drain passage 17 through a direction switching valve 15 including a four-port three-position electromagnetic switching valve as described later. It is like that. The leading end of the drain passage 17 is open to the oil pan 18.
[0035]
The upstream side of the hydraulic pressure introduction passage 16 branches into a low pressure side hydraulic pressure introduction passage 21 and a high pressure side hydraulic pressure introduction passage 22. The low pressure side hydraulic pressure introduction passage 21 lubricates an internal combustion engine as a low pressure hydraulic pressure source. Communicates with the oil gallery. In other words, a relatively low pressure oil is supplied by a lubricating oil pump (not shown) of the internal combustion engine. The low pressure side hydraulic pressure introduction passage 21 is provided with a check valve 23 for preventing the backflow of hydraulic oil from the high pressure side hydraulic pressure introduction passage 22.
[0036]
On the other hand, an electric high pressure hydraulic pump 25 and a hydraulic accumulator 26 are connected to the high pressure side hydraulic pressure introduction passage 22 as a high pressure hydraulic pressure source. The electric high-pressure hydraulic pump 25 is driven by an electric motor 25a. The electric high-pressure hydraulic pump 25 is ON / OFF controlled based on the detection of a hydraulic sensor 27 provided in the hydraulic accumulator 26, and the hydraulic accumulator. A high hydraulic pressure is always accumulated in 26. Note that a check valve 28 is interposed between the high pressure hydraulic pump 25 and the hydraulic accumulator 26 to prevent backflow of hydraulic oil. Further, an on-off valve 29 that turns on and off the supply of high-pressure hydraulic pressure from the accumulator 26 is interposed downstream of the hydraulic accumulator 26.
[0037]
Next, the operation of the hydraulic device 6 will be described.
[0038]
FIG. 9 shows a control state in the case where the above-described high compression ratio state is maintained. That is, the on-off valve 29 is closed, and the direction switching valve 15 blocks the first hydraulic chamber 12 and the drain passage 17 and connects the second hydraulic chamber 13 to the hydraulic pressure introduction passage 16. Accordingly, the relatively low pressure hydraulic oil of the engine lubrication system is introduced into the second hydraulic chamber 13 through the low pressure side hydraulic pressure introduction passage 21 so as to compensate for the leakage amount of the hydraulic actuator 5, and the rotational position of the hydraulic actuator 5 is maintained. Is done. During this time, the electric high pressure hydraulic pump 25 is basically stopped.
[0039]
If the position is simply held, the rotational force (moment load) acting on the control shaft 62 is small as described above, and therefore, the low pressure oil of the lubricating system is sufficient. In the case of a vane actuator, leakage at a high pressure becomes a problem. However, if the order is several kg / cm ² like the oil pressure of lubricating oil, the amount of leakage becomes very small in proportion to the oil pressure. In addition, the oil pump for engine lubrication discharges oil abundantly, and even with the same amount of leakage, the efficiency is several times better than the electric high pressure hydraulic pump 25, so the impact on energy consumption is extremely small. Become. In this way, it is almost unnecessary to drive the electric high-pressure hydraulic pump 25 in order to maintain the high compression ratio state that is frequently used, and the fuel efficiency improvement effect by the variable compression ratio control is not diminished.
[0040]
Next, FIG. 10 shows the state of the direction switching valve 15 and the like when switching from the high compression ratio state to the low compression ratio. In the case of switching to this low compression ratio, a high switching speed is required to avoid the occurrence of knocking. Accordingly, in this case, the direction switching valve 15 is switched to the illustrated position for connecting the first hydraulic chamber 12 to the hydraulic pressure introduction passage 16 and the second hydraulic chamber 13 to the drain passage 17, and the on-off valve 29. Is switched to open, and high-pressure hydraulic oil is quickly introduced from the hydraulic accumulator 26 to the first hydraulic chamber 12. As a result, the hydraulic actuator 5 immediately rotates and is switched to the low compression ratio state. In this case, the check valve 23 prevents the high-pressure hydraulic oil from flowing out to the oil gallery. After switching to the low compression ratio, it is possible to keep introducing the high-pressure hydraulic oil into the first hydraulic chamber 12 while maintaining the state shown in the figure. The energy consumed can be minimized by returning to the neutral position shown in FIG. 9 and turning off the on-off valve 29 and holding the position of the hydraulic actuator 5 with the hydraulic pressure of the lubricating system, which is a low-pressure hydraulic source.
[0041]
FIG. 11 shows the state of the direction switching valve 15 and the like when switching from the low compression ratio state to the high compression ratio. In switching to the high compression ratio side, an appropriate switching speed required for improving fuel efficiency is required. The direction switching valve 15 is switched to the state shown in the figure, and the on-off valve 29 is opened to introduce high-pressure hydraulic oil into the second hydraulic chamber 13 of the hydraulic actuator 5. In this case as well, the check valve 23 prevents the high pressure hydraulic oil from flowing out to the oil gallery. After switching to the high compression ratio, in order to minimize energy consumption, the direction switching valve 15 is returned to the neutral position and the on-off valve 29 is closed as described above, and the hydraulic actuator 5 is driven by the oil pressure of the lubricating system. Hold.
[0042]
Note that the rotational force (moment load) acting on the control shaft 62 changes cyclically with the combustion explosion of each cylinder, so that the hydraulic pressure of the second hydraulic chamber 13 on the load side when the position of the hydraulic actuator 5 is held is maintained. Even fluctuate cyclically. Therefore, even if the oil pressure of the lubrication system of the internal combustion engine is low, the working oil is introduced through the low pressure side oil pressure introduction passage 21 by the action of the check valve 23 during the period when the oil pressure in the second hydraulic chamber 13 is lowered. Or replenishment is performed. That is, even at an intermediate rotational position, the rotational position can be held by the low pressure hydraulic source.
[0043]
FIG. 12 shows the configuration of an intake / exhaust system including the turbocharger 101. The turbocharger 101 has a configuration in which a turbine 103 located in the exhaust passage 102 of the internal combustion engine and a compressor 105 located in the intake passage 104 are coaxially connected, and controls the supercharging pressure in accordance with operating conditions. In addition, an exhaust bypass valve 106 for bypassing a part of the exhaust from the upstream side of the turbine 103 is provided. Further, a supercharging pressure sensor 107 that detects a supercharging pressure is disposed on the downstream side of the compressor 105 in the intake passage 104, and the knock sensor 3 is disposed in the cylinder block 50 as described above. 108 is an intake valve and 109 is an exhaust valve. In this example, a variable valve mechanism (not shown) capable of variably controlling the closing timing (IVC) of the intake valve 108 is provided, and the actuator 110 of this variable valve mechanism is simultaneously controlled by the engine control unit 4. .
[0044]
FIG. 13 shows an example of a specific flowchart for realizing the compression ratio control described above. After the above-described map of the target compression ratio ε is read (step 1), engine operating conditions (rotation speed, load, etc.) are detected (step 2). Next, the hydraulic pressure of the hydraulic accumulator 26 is determined (step 3). If the hydraulic pressure is insufficient, the electric high pressure hydraulic pump 25 is operated (step 4). Next, a target compression ratio ε is set corresponding to the engine operating condition (step 5), and it is determined whether the current compression ratio ε matches the target value (step 6). If the current compression ratio ε matches the target value, the direction switching valve (abbreviated as OCV in the flowchart) 15 is set to the neutral position (step 7), and the on-off valve 29 is closed (step 8). Accordingly, as described above, the holding by the low pressure hydraulic power source is performed. On the other hand, if the current compression ratio ε does not match the target value, it is determined whether the operation is to be performed to the low compression ratio side or the high compression ratio side (step 9), and the direction is switched in accordance with each direction. The valve 15 is switched (steps 10 and 11). Then, the on-off valve 29 is opened (step 12). As a result, the high-pressure hydraulic oil is supplied from the hydraulic accumulator 26, and the compression ratio ε changes quickly. After the compression ratio ε coincides with the target value due to the operation of the hydraulic actuator 5, the direction switching valve 15 is set to the neutral position (step 7), and the on-off valve 29 is closed (step 8). It will be held by.
[Brief description of the drawings]
FIG. 1 is an explanatory diagram showing a configuration of an embodiment of a compression ratio control device for an internal combustion engine according to the present invention.
FIG. 2 is a cross-sectional view showing a configuration of a variable compression ratio mechanism.
FIG. 3 is a side view showing configurations of a control shaft and a hydraulic actuator.
FIG. 4 is an explanatory diagram showing an example of a link configuration.
FIG. 5 is a characteristic diagram showing characteristics of a target compression ratio ε with respect to operating conditions.
FIG. 6 is a characteristic diagram showing a change in moment load depending on an angle θ formed by an eccentric direction and a load vector.
FIG. 7 is an explanatory diagram showing a relationship between a posture of a control shaft and a load vector during high compression ratio control.
FIG. 8 is an explanatory diagram showing the relationship between the posture of the control shaft and the load vector during low compression ratio control.
FIG. 9 is a hydraulic circuit diagram showing the configuration of the hydraulic device in a state where it is held in a high compression ratio state.
FIG. 10 is a hydraulic circuit diagram showing a state when switching to a low compression ratio side.
FIG. 11 is a hydraulic circuit diagram showing a state when switching to a high compression ratio side.
FIG. 12 is an explanatory diagram showing a configuration of an intake / exhaust system including a turbocharger.
FIG. 13 is a flowchart showing a flow of compression ratio control.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 5 ... Hydraulic actuator 15 ... Direction switching valve 25 ... Electric high pressure hydraulic pump 26 ... Hydraulic accumulator 29 ... On-off valve 60 ... Control link 62 ... Control shaft 62a ... Eccentric cam part

Claims (5)

A first link connected to a piston that reciprocates in a cylinder via a piston pin, and a second link that is swingably connected to the first link and is rotatably attached to a crankpin of a crankshaft. A third link whose one end is swingably connected to the second link, and an eccentric cam portion rotatably provided on the cylinder block and rotatably supporting the other end of the third link In a compression ratio control device for an internal combustion engine, comprising: a control shaft provided with an actuator for controlling the rotational position of the control shaft, and changing the engine compression ratio according to the rotational position of the control shaft.
The rotation angle range of the control shaft from the rotational position where the compression ratio is the highest value to the rotational position where the compression ratio is the lowest value is approximately 180 °,
When the compression ratio is controlled to the highest value, a plane passing through the center line of the control shaft and the center line of the eccentric cam portion becomes a load vector of a load acting on the eccentric cam portion from the third link. As opposed to being almost parallel ,
A compression ratio control device for an internal combustion engine, wherein the plane is substantially parallel to the load vector when the compression ratio is controlled to the lowest value . When the compression ratio is controlled to the highest value, the plane is not completely parallel to the load vector, and a moment for rotating the control shaft to the low compression ratio side is generated by the load vector. The compression ratio control device for an internal combustion engine according to claim 1 , wherein the compression ratio control device is configured as follows. The compression ratio control device for an internal combustion engine according to claim 2 , wherein a rotation angle range of the control shaft is slightly smaller than 180 °. A hydraulic accumulator is used as the actuator, and when the engine compression ratio is maintained at a constant compression ratio, a leakage amount is replenished to the hydraulic actuator by hydraulic pressure supplied from a lubricating oil pump of the internal combustion engine. Item 4. The compression ratio control device for an internal combustion engine according to any one of Items 1 to 3 . 5. The internal combustion engine according to claim 4 , further comprising a high-pressure hydraulic source different from the lubricating oil pump, wherein hydraulic oil is supplied from the high-pressure hydraulic source to the hydraulic actuator when the engine compression ratio is changed. Engine compression ratio control device.

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